Automatic transmission

ABSTRACT

Provided is an automatic transmission capable of reducing friction loss. The automatic transmission is provided with an input shaft ( 2 ) coupled with a sun gear (Sa), an output member ( 3 ), a first to third planetary gear mechanisms ( 4, 5, 6 ), a first engagement mechanism (C 1 ) coupling the input shaft ( 2 ) with a ring gear (Rc) releasably, a second engagement mechanism (C 2 ) coupling a carrier (Ca) with a carrier (Cc) releasably, a third engagement mechanism (C 3 ) coupling the carrier (Ca) with a second coupling body (Rb-Sc) releasably, a fourth engagement mechanism (B 1 ) fixing a first coupling body (Ra-Cb) to a transmission case ( 1 ) releasably, a fifth engagement mechanism (B 2 ) fixing a sun gear (Sb) to the transmission case ( 1 ) releasably, and a sixth engagement mechanism (B 3 ) fixing the ring gear (Rc) to the transmission case ( 1 ) releasably.

CROSS-REFERENCE TO RELATED APPLICATIONS

This application is a Divisional of U.S. patent Application Ser. No.12/771,112, filed Apr. 30, 2010, and which is based upon and claims thebenefit of priority from the prior Japanese Patent Application No.2009-123832, filed on May 22, 2009 and Japanese Patent Application No.2009-142142, filed on Jun. 15, 2009, the entire contents of which areincorporated herein by reference.

BACKGROUND OF THE INVENTION 1. Field of the Invention

The present invention relates to an automatic transmission which changesrotations of an input shaft into multiple gear speeds transmitted to anoutput member via a plurality of planetary gear mechanisms disposed in atransmission case.

2. Description of the Related Art

Conventionally, there has been known an automatic transmission (forexample, refer to Japanese Patent Laid-Open No. 2005-273768) whichperforms 8 forward gear speed changes via a first planetary gear forinputting, two planetary gears for gear change and six engagementmechanisms.

The first planetary gear in the automatic transmission disclosed inJapanese Patent Laid-Open No. 2005-273768 is of a double-pinionplanetary gear mechanism comprised of a first sun gear, a first ringgear, and a first carrier pivotally supporting a pair of first pinions,which are intermeshed with each other with either one intermeshed withthe first sun gear and the other one intermeshed with the first ringgear, in such a way that the pair of first pinions can rotate andrevolve freely.

In the first planetary gear, the sun gear is equivalent to a fixingelement fixed to a transmission case, the first carrier is equivalent toan input element coupled with an input shaft, and the first ring gear isequivalent to an output element configured to decelerate a rotationspeed of the first carrier equivalent to the input element and outputthe decelerated rotation speed.

The two planetary gears for gear change are of a Ravigneaux planetarygear mechanism comprised of a second sun gear, a third sun gear, asecond ring gear integral with a third ring gear, and a second carrierpivotally supporting a pair of second pinions, which are intermeshedwith each other with either one intermeshed with the second sun gear andthe second ring gear and the other one intermeshed with the third sungear, in such a way that the pair of second pinions can rotate andrevolve freely.

In the Ravigneaux planetary gear mechanism, the second sun gear isequivalent to a first rotation element, the second carrier integral withthe third carrier is equivalent to a second rotation element, the secondring gear integral with the third ring gear is equivalent to a thirdrotation element, and the third sun gear is equivalent to a fourthrotation element in the order of distances relative to gear ratios in avelocity diagram.

The engagement mechanisms are comprised of a first engagement mechanismreleasably coupling the first ring gear equivalent to the output elementof the first planetary gear with the fourth rotation element composed ofthe third sun gear, a second engagement mechanism releasably couplingthe input shaft with the second rotation element composed of the secondcarrier, a third engagement mechanism releasably coupling the first ringgear equivalent to the output element with the first rotation elementcomposed of the second sun gear, a fourth engagement mechanismreleasably coupling the first carrier equivalent to the input elementwith the first rotation element composed of the second sun gear, a fifthengagement mechanism releasably fixing the first rotation elementcomposed of the second sun gear to the transmission case, and a sixthengagement mechanism releasably fixing the second rotation elementcomposed of the second carrier to the transmission case.

According to the configurations mentioned above, a first gear speed isestablished when the first engagement mechanism and the sixth engagementmechanism are engaged; a second gear speed is established when the firstengagement mechanism and the fifth engagement mechanism are engaged; athird gear speed is established when the first engagement mechanism andthe third engagement mechanism are engaged; a fourth gear speed isestablished when the first engagement mechanism and the fourthengagement mechanism are engaged.

A fifth gear speed is established when the first engagement mechanismand the second engagement mechanism are engaged; a sixth gear speed isestablished when the second engagement mechanism and the fourthengagement mechanism are engaged; a seventh gear speed is establishedwhen the second engagement mechanism and the third engagement mechanismare engaged; an eighth gear speed is established when the secondengagement mechanism and the fifth engagement mechanism are engaged. Inaddition, if the first gear speed or the eighth gear speed is excluded,it is possible to establish 7 forward gear speeds.

In the conventional transmission, the engagement mechanisms to beengaged in each gear speed are two. Thereby, the dragging among theother four freed engagement mechanisms will make friction loss becomegreater, deteriorating the efficiency of the transmission.

SUMMARY OF THE INVENTION

The present invention has been accomplished in view of theaforementioned problems, and it is therefore an object of the presentinvention to provide an automatic transmission capable of reducingfriction loss.

A first aspect of the present invention provides an automatictransmission which changes rotations of an input shaft into multiplegear speeds transmitted to an output member via a plurality of planetarygear mechanisms disposed in a transmission case, wherein the pluralityof planetary gear mechanisms includes three planetary gear mechanisms ofa first planetary gear mechanism to a third planetary gear mechanism,three elements comprised of a sun gear, a carrier and a ring gear of thefirst planetary gear mechanism are set as a first element, a secondelement and a third element, respectively, in the order of distancesrelative to gear ratios in a velocity diagram, three elements comprisedof a sun gear, a carrier and a ring gear of the second planetary gearmechanism are set as a fourth element, a fifth element and a sixthelement, respectively, in the order of distances relative to gear ratiosin the velocity diagram, three elements comprised of a sun gear, acarrier and a ring gear of the third planetary gear mechanism are set asa seventh element, an eighth element and a ninth element, respectively,in the order of distances relative to gear ratios in the velocitydiagram, the first element is coupled with the input shaft, the seventhelement is coupled with the output member, the third element and thefifth element are coupled to form a first coupling body, the sixthelement and the ninth element are coupled to form a second couplingbody, a first engagement mechanism couples the input shaft with theeighth element releasably, a second engagement mechanism couples thesecond element with the seventh element releasably, a third engagementmechanism couples the second element with the second coupling bodyreleasably, a fourth engagement mechanism fixes the first coupling bodyto the transmission case releasably, a fifth engagement mechanism fixesthe fourth element to the transmission case releasably, a sixthengagement mechanism fixes the eighth element to the transmission casereleasably, and the third planetary gear mechanism is of a double-pinionplanetary gear mechanism comprised of the sun gear, the ring gear, andthe carrier pivotally supporting a pair of pinions, which areintermeshed with each other with either one intermeshed with the sungear and the other one intermeshed with the ring gear, in such a waythat the pair of pinions can rotate and revolve freely.

As to be made clear by an embodiment to be described hereinafter,according to the first aspect of the present invention the automatictransmission can perform 7 forward gear speed changes or even more, andin each gear speed change, three of the six engagement mechanisms fromthe first to the sixth engagement mechanism are engaged. Thereby, ineach gear speed change, the freed engagement mechanisms are three. Incomparison with the conventional transmission in which the freedengagement mechanisms are four, the friction loss caused by the freedengagement mechanisms is reduced, and consequently, the efficiency ofthe automatic transmission is improved.

In the first aspect of the present invention, it is preferable that thesixth engagement mechanism is a one-way clutch or a two-way clutch.According to the mentioned configuration, it is possible to bettercontrol the gear changes between the first gear speed and the secondgear speed in comparison with the case where the sixth engagementmechanism is formed from a wet multi-plate clutch only.

In the first aspect of the present invention, it is preferable that thesecond engagement mechanism is an intermeshing mechanism. As to be madeclear by an embodiment to be described hereinafter, according to thementioned configuration, the second intermeshing mechanism which isengaged only in a range of low forward gear speeds and freed only in arange of high forward gear speeds functions as a mechanism avoiding thefriction loss. Thereby, the friction loss can be further inhibited inthe range of high forward gear speeds, and consequently, the efficiencyof the automatic transmission is further improved.

In the first aspect of the present invention, it is preferable that thethird engagement mechanism is disposed at an outer position of thesecond engagement mechanism in the radial direction thereof and has atleast a part overlapped with the second engagement mechanism in theaxial direction of the input shaft. According to the mentionedconfiguration, the length of the shaft of the automatic transmission canbe made shorter.

A second aspect of the present invention provides an automatictransmission which changes rotations of an input shaft into multiplegear speeds transmitted to an output member via a plurality of planetarygear mechanisms disposed in a transmission case , wherein the pluralityof planetary gear mechanisms includes three planetary gear mechanisms ofa first planetary gear mechanism to a third planetary gear mechanism,the third planetary gear mechanism is of a single-pinion planetary gearmechanism comprised of the sun gear, the ring gear, and the carrierpivotally supporting a pinion intermeshed with the sun gear and the ringgear in such a way that the pinion can rotate and revolve freely, threeelements comprised of a sun gear, a carrier and a ring gear of the firstplanetary gear mechanism are set as a first element, a second elementand a third element, respectively, in the order of distances relative togear ratios in a velocity diagram, three elements comprised of a sungear, a carrier and a ring gear of the second planetary gear mechanismare set as a fourth element, a fifth element and a sixth element,respectively, in the order of distances relative to gear ratios in thevelocity diagram, three elements comprised of a sun gear, a carrier anda ring gear of the third planetary gear mechanism are set as a seventhelement, an eighth element and a ninth element, respectively, in theorder of distances relative to gear ratios in the velocity diagram, aparallel gear is disposed to be adjacent to the parallel gear andintermeshed with the pinion of the third planetary gear mechanism, thefirst element is coupled with the input shaft, the seventh element iscoupled with the output member, the third element and the fifth elementare coupled to form a first coupling body, the sixth element and theninth element are coupled to form a second coupling body, a firstengagement mechanism couples the input shaft with the eighth elementreleasably, a second engagement mechanism couples the second elementwith the seventh element releasably, a third engagement mechanismcouples the second element with the second coupling body releasably, afourth engagement mechanism fixes the first coupling body to thetransmission case releasably, a fifth engagement mechanism fixes thefourth element to the transmission case releasably, and a sixthengagement mechanism fixes the eighth element to the transmission casereleasably.

As to be made clear by an embodiment to be described hereinafter,according to the second aspect of the present invention the automatictransmission can perform 7 forward gear speed changes or even more, andin each gear speed change, three of the six engagement mechanisms fromthe first to the sixth engagement mechanism are engaged. Thereby, ineach gear speed change, the freed engagement mechanisms are three. Incomparison with the conventional transmission in which the freedengagement mechanisms are four, the friction loss caused by the freedengagement mechanisms is reduced, and consequently, the efficiency ofthe automatic transmission is improved.

In the second aspect of the present invention, in addition to the thirdplanetary gear mechanism, it is possible to configure the firstplanetary gear mechanism and the second planetary gear mechanism as asingle-pinion planetary gear mechanism comprised of the sun gear, thering gear, and the carrier pivotally supporting a pinion intermeshedwith the sun gear and the ring gear in such a way that the pinion canrotate and revolve freely.

According to the mentioned configuration, by configuring all the threeplanetary gear mechanisms into the single-pinion planetary mechanism, incomparison with the case of the double-pinion planetary mechanism, thenumber of intermeshing times can be reduced, and consequently, theefficiency of the automatic transmission is improved.

In the second aspect of the present invention, similar to the firstaspect, it is preferable that the sixth engagement mechanism is aone-way clutch or a two-way clutch. According to the mentionedconfiguration, it is possible to better control the gear changes betweenthe first gear speed and the second gear speed in comparison with thecase where the sixth engagement mechanism is formed from a wetmulti-plate clutch only.

In the second aspect of the present invention, similar to the firstaspect, it is preferable that the second engagement mechanism is anintermeshing mechanism. As to be made clear by an embodiment to bedescribed hereinafter, according to the mentioned configuration, thesecond engagement mechanism which is engaged only in a range of lowforward gear speeds and freed only in a range of high forward gearspeeds functions as a mechanism avoiding the friction loss. Thereby, thefriction loss can be further inhibited in the range of high forward gearspeeds, and consequently, the efficiency of the automatic transmissionis further improved.

In the second aspect of the present invention, when the automatictransmission is used in a FR layout vehicle, for the purpose ofminiaturization, it is desired that the output member is configured tobe a shaft as the output shaft, disposed coaxially with the input shaftand coupled with a propeller shaft or the like. However, for example, ifthe parallel gear is disposed between the fifth engagement mechanism andthe sixth engagement mechanism, it is impossible to dispose the outputshaft, namely the output member, coaxially with the input shaft due tothe obstacle from the transmission case.

Therefore, it is necessary to use a gear as the output member, dispose acounter shaft in parallel to the input shaft, a follower gearintermeshed with the gear equivalent to the output member to the countershaft. Thereby, the driving force is transmitted from the counter shaftto the rear wheels via the propeller shaft or the like. However, in thiscase, it is necessary to dispose the counter shaft therein, which makesthe automatic transmission become large-sized unfavorably.

In this regarding, in the present invention, it is favorable that thefourth to the sixth engagement mechanisms are disposed closer to adriving source for driving the input shaft to rotate than to theparallel gear, and the output member is an output shaft disposedcoaxially with the input shaft. According to the mentionedconfiguration, the output member is configured as a shaft and isdisposed coaxially with the input shaft, and can be coupled with theparallel gear without being obstructed by the transmission case, whichmakes it possible to miniaturize the transmission.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a skeleton diagram of an automatic transmission according to afirst embodiment of the present invention.

FIG. 2 is a velocity diagram for the automatic transmission according tothe first embodiment of the present invention.

FIG. 3 is an explanatory diagram illustrating engagement states ofengagement mechanisms in each gear speed of the automatic transmissionaccording to the first embodiment.

FIG. 4 is a skeleton diagram of an automatic transmission according to asecond embodiment of the present invention.

FIG. 5 is a skeleton diagram of an automatic transmission according to athird embodiment of the present invention.

FIG. 6 is a skeleton diagram of an automatic transmission according to afourth embodiment of the present invention.

FIG. 7 is a velocity diagram for planetary gears according to the fourthembodiment of the present invention.

FIG. 8 is an explanatory diagram illustrating engagement states ofengagement mechanisms in each gear speed of the automatic transmissionaccording to the fourth embodiment.

FIG. 9 is a skeleton diagram of an automatic transmission according to afifth embodiment of the present invention.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS

An automatic transmission according to a first embodiment of the presentinvention is illustrated in FIG. 1. The automatic transmission of thefirst embodiment is provided with a transmission case 1, an input shaft2 and an output member 3. The input shaft 2 is pivotally supportedinside the transmission case 1 and coupled with a driving source such asan engine (not shown). The output member 3 is comprised of output gearsdisposed concentrically with the input shaft 2. Rotations of the outputmember 3 are transmitted to driving wheels disposed at both sides of avehicle via a differential gear (not shown).

Further, a first planetary gear mechanism 4, a second planetary gearmechanism 5 and a third planetary gear mechanism 6 are disposedconcentrically with the input shaft 2 inside the transmission case 1.The first planetary gear mechanism 4 is a single-pinion planetary gearmechanism comprised of a sun gear Sa, a ring gear Ra, and a carrier Capivotally supporting a pinion Pa intermeshed with the sun gear Sa andthe ring gear Ra in such a way that the pinion Pa can rotate and revolvefreely.

Referring to the top section of a velocity diagram (a diagramillustrating rotation velocities of three elements of the sun gear, thecarrier and the ring gear by straight lines) for the first planetarygear mechanism 4 in FIG. 2, if the three elements composed of the sungear Sa, the carrier Ca and the ring gear Ra of the first planetary gearmechanism 4 are arranged from the left side in the order of distancesrelative to gear ratios in the velocity diagram, they are equivalent tothe first element, the second element and the third element,respectively.

Herein, when the gear ratio (number of teeth of the ring gear/number ofteeth of the sun gear) of the first planetary gear mechanism 4 issupposed to be “i”, the ratio of a distance between the sun gear Sa andthe carrier Ca to a distance between the carrier Ca and the ring gear Rais set to i:1. In the velocity diagram, the lower horizontal line andthe upper horizontal line indicate a rotational speed of “0” and arotational speed of “1” (equal to that of the input shaft 2),respectively.

The second planetary gear mechanism 5 is of a single-pinion planetarygear mechanism comprised of a sun gear Sb, a ring gear Rb and a carrierCb pivotally supporting a pinion Pb intermeshed with the sun gear Sb andthe ring gear Rb in such a way that the pinion Pb can rotate and revolvefreely.

Referring to the middle section of a velocity diagram for the secondplanetary gear mechanism 5 in FIG. 2, if the three elements composed ofthe sun gear Sb, the carrier Cb and the ring gear Rb of the secondplanetary gear mechanism 5 are arranged from the left side in the orderof distances relative to gear ratios in the velocity diagram, they areequivalent to the fourth element, the fifth element and the sixthelement, respectively. When the gear ratio of the second planetary gearmechanism 5 is supposed to be “j”, the ratio of a distance between thesun gear Sb and the carrier Cb to a distance between the carrier Cb andthe ring gear Rb is set to j:1.

The third planetary gear mechanism 6 is of a double-pinion planetarygear mechanism comprised of a sun gear Sc, a ring gear Rc, and a carrierCc pivotally supporting a pair of pinions Pc and Pc′, which areintermeshed with each other with either one intermeshed with the sungear Sc and the other one intermeshed with the ring gear Rc, in such away that the pair of pinions Pc and Pc′ can rotate and revolve freely.

Referring to the bottom section of a velocity diagram for the thirdplanetary gear mechanism 6 in FIG. 2, if the three elements composed ofthe carrier Cc, the ring gear Rc and the sun gear Sc of the thirdplanetary gear mechanism 6 are arranged from the left side in the orderof distances relative to gear ratios in the velocity diagram, they areequivalent to the seventh element, the eighth element and the ninthelement, respectively. When the gear ratio of the third planetary gearmechanism 6 is supposed to be “k”, the ratio of a distance between thesun gear Sc and the carrier Cc to a distance between the carrier Cc andthe ring gear Rc is set to k:1.

The sun gear Sa (the first element) of the first planetary gearmechanism 4 is coupled with the input shaft 2. The carrier Cc (theseventh element) of the third planetary gear mechanism 6 is coupled withthe output member 3.

The ring gear Ra (the third element) of the first planetary gearmechanism 4 and the carrier Cb (the fifth element) of the secondplanetary gear mechanism 5 are coupled with each other to form a firstcoupling body Ra-Cb. The ring gear Rb (the sixth element) of the secondplanetary gear mechanism 5 and the sun gear Sc (the ninth element) ofthe third planetary gear mechanism 6 are coupled with each other to forma second coupling body Rb-Sc.

In the automatic transmission of the first embodiment, a total number of7 rotation bodies are constituted in the three planetary gear mechanisms4, 5 and 6, specifically, the sun gear Sa (the first element) and thecarrier Ca (the second element) of the first planetary gear mechanism 4,the first coupling body Ra-Cb (the third-fifth elements), the sun gearSb (the fourth element) of the second planetary gear mechanism 5, thesecond coupling body Rb-Sc (the sixth-ninth elements), the carrier Cc(the seventh element) and the ring gear Rc (the eighth element) of thethird planetary gear mechanism 6.

The automatic transmission of the first embodiment is provided withengagement mechanisms comprised of wet multi-plate clutches,specifically, a first clutch C1 equivalent to a first engagementmechanism coupled with the input shaft 2 and the ring gear Rc (theeighth element) of the third planetary gear mechanism 6 releasably, asecond clutch C2 equivalent to a second engagement mechanism coupledwith the carrier Ca (the second element) of the first planetary gearmechanism 4 and the carrier Cc (the seventh element) of the thirdplanetary gear mechanism 6 releasably, and a third clutch C3 equivalentto a third engagement mechanism coupled with the carrier Ca (the secondelement) of the first planetary gear mechanism 4 and the second couplingbody Rb-Sc (the sixth-ninth elements) releasably.

The clutch C3 equivalent to the third engagement mechanism is disposedat an outer position of the second clutch C2 equivalent to the secondengagement mechanism in the radial direction thereof and overlapped withthe second clutch C2 in the axial direction of the input shaft 2 toshorten the shaft length of the automatic transmission.

The automatic transmission of the first embodiment is provided withengagement mechanisms comprised of wet multi-plate brakes, specifically,a first brake B1 equivalent to a fourth engagement mechanism fixing thefirst coupling body Ra-Cb (the third-fifth elements) to the transmissioncase 1 releasably, a second brake B2 equivalent to a fifth engagementmechanism fixing the sun gear Sb of the second planetary gear mechanism5 to the transmission case 1 releasably, and a third brake B3 fixing thering gear Rc (the eighth element) of the third planetary gear mechanism6 to the transmission case 1 releasably.

A 1-way clutch F1 is disposed in parallel with the third brake B3 insidethe transmission case 1, allowing the ring gear Rc (the eighth element)of the third planetary gear mechanism 6 to rotate positively (forwardrotation) and preventing it from rotating negatively (reverse rotation).

The third brake B3 and the 1-way clutch F1 in the automatic transmissionof the first embodiment constitute a sixth engagement mechanism of thepresent invention.

When the second clutch C2 (the second engagement mechanism) and thesecond brake B2 (the fifth engagement mechanism) are engaged in theautomatic transmission of the first embodiment, the carrier Ca (thesecond element) of the first planetary gear mechanism 4 and the carrierCc (the seventh element) of the third planetary gear mechanism 6 rotateat the same rotation speed, the rotation speed of the sun gear Sb (thefourth element) of the second planetary gear mechanism 5 becomes equalto “0”, and the rotation speed of the ring gear Rc (the eighth element)of the third planetary gear mechanism 6 becomes equal to “0” due to thefunction of the 1-way clutch F1. Thereby, the velocity line of the threeplanetary gear mechanisms 4, 5 and 6 becomes a line denoted by “1st” inFIG. 2, and the first gear speed is established.

At this time, the third brake B3 is released, however, since therotation speed of the ring gear Rc (the eighth element) of the thirdplanetary gear mechanism 6 becomes equal to “0” due to the function ofthe 1-way clutch F1, no friction loss will be happened in the thirdbrake B3. Moreover, the disposition of the 1-way clutch F1 makes itunnecessary to supply pressured oils to the third brake B3 and to stopsupplying pressured oils thereto when the gear change is made betweenthe first gear speed and the second gear speed, which improves thecontrol on gear change between the first gear speed and the second gearspeed.

In addition to the second clutch C2 (the second engagement mechanism)and the second brake B2 (the fifth engagement mechanism), when the thirdbrake B3 is further engaged, the first gear speed is established withthe engine braking in action.

When the second engagement mechanism C2, the fourth engagement mechanismB1 and the fifth engagement mechanism B2 are engaged, both the rotationspeed of the first coupling body Ra-Cb (the third-fifth elements) andthe rotation speed of the sun gear Sb (the fourth element) of the secondplanetary gear mechanism 5 become equal to “0”, the three elements ofthe second planetary gear mechanism 5 are locked in a state whererelative rotations are impossible, therefore, the rotation speed of thesecond coupling body Rb-Sc (the sixth-ninth elements) also becomes equalto “0”.

Thereby, the carrier Ca (the second element) of the first planetary gearmechanism 4 and the carrier Cc (the seventh element) of the thirdplanetary gear mechanism 6 rotate at the same rotation speed, thevelocity line of the three planetary gear mechanisms 4, 5 and 6 becomesa line denoted by “2nd” in FIG. 2, and the second gear speed isestablished.

When the second engagement mechanism C2, the third engagement mechanismC3 and the fifth engagement mechanism B2 are engaged, the carrier Ca(the second element) of the first planetary gear mechanism 4, the secondcoupling body Rb-Sc (the sixth-ninth elements) and the carrier Cc (theseventh element) of the third planetary gear mechanism 6 rotate at thesame rotation speed, the three elements, namely the sun gear Sc, thecarrier Cc and the ring gear Rc of the third planetary gear mechanism 6are locked in a state where relative rotations are impossible. Thereby,the velocity line of the three planetary gear mechanisms 4, 5 and 6becomes a line denoted by “3rd” in FIG. 2, and the third gear speed isestablished.

When the first engagement mechanism C1, the second engagement mechanismC2 and the fifth engagement mechanism B2 are engaged, the rotation speedof the sun gear Sa (the first element) of the first planetary gearmechanism 4 and the rotation speed of the ring gear Rc (the eighthelement) of the third planetary gear mechanism 6 become equal to “1”,and the carrier Ca (the second element) of the first planetary gearmechanism 4 and the carrier Cc (the seventh element) of the thirdplanetary gear mechanism 6 rotate at the same rotation speed. Thereby,the velocity line of the three planetary gear mechanisms 4, 5 and 6becomes a line denoted by “4th” in FIG. 2, and the fourth gear speed isestablished.

When the first engagement mechanism C1, the second engagement mechanismC2 and the third engagement mechanism C3 are engaged, the rotation speedof the sun gear Sa (the first element) of the first planetary gearmechanism 4 and the rotation speed of the ring gear Rc (the eighthelement) of the third planetary gear mechanism 6 become equal to “1”,the three elements, namely the sun gear Sc, the carrier Cc and the ringgear Rc of the third planetary gear mechanism 6 are locked in a statewhere relative rotations are impossible. Thereby, the fifth gear speedis established at “1” which is also the rotation speed of the carrier Ccof the third planetary gear mechanism 6 coupled with the output member3.

When the first engagement mechanism C1, the third engagement mechanismC3 and the fifth engagement mechanism B2 are engaged, the rotation speedof the sun gear Sa (the first element) of the first planetary gearmechanism 4 and the rotation speed of the ring gear Rc (the eighthelement) of the third planetary gear mechanism 6 become equal to “1”,the rotation speed of the sun gear Sb (the fourth element) of the secondplanetary gear mechanism 5 becomes equal to “0”, and the carrier Ca (thesecond element) of the first planetary gear mechanism 4 and the secondcoupling body Rb-Sc (the sixth-ninth elements) rotate at the samerotation speed. Thereby, the velocity line of the three planetary gearmechanisms 4, 5 and 6 becomes a line denoted by “6th” in FIG. 2, and thesixth gear speed is established.

When the first engagement mechanism C1, the third engagement mechanismC3 and the fourth engagement mechanism B1 are engaged, the rotationspeed of the sun gear Sa (the first element) of the first planetary gearmechanism 4 and the rotation speed of the ring gear Rc (the eighthelement) of the third planetary gear mechanism 6 become equal to “1”,the rotation speed of the first coupling body Ra-Cb (the third-fifthelements) becomes equal to “0”, and the carrier Ca (the second element)of the first planetary gear mechanism 4 and the second coupling bodyRb-Sc (the sixth-ninth elements) rotate at the same rotation speed.Thereby, the velocity line of the three planetary gear mechanisms 4, 5and 6 becomes a line denoted by “7th” in FIG. 2, and the seventh gearspeed is established.

When the first engagement mechanism C1, the fourth engagement mechanismB1 and the fifth engagement mechanism B2 are engaged, the rotation speedof the sun gear Sa (the first element) of the first planetary gearmechanism 4 and the rotation speed of the ring gear Rc (the eighthelement) of the third planetary gear mechanism 6 become equal to “1”,the rotation speed of the sun gear Sb (the fourth element) of the secondplanetary gear mechanism 5 and the rotation speed of the first couplingbody Ra-Cb (the third-fifth elements) become equal to “0”.

Thereby, the three elements, namely the sun gear Sb, the carrier Cb andthe ring gear Rb of the second planetary gear mechanism 5 are locked ina state where relative rotations are impossible, the rotation speed ofthe second coupling body Rb-Sc (the sixth-ninth elements) becomes equalto “0”. Thereby, the velocity line of the three planetary gearmechanisms 4, 5 and 6 becomes a line denoted by “8th” in FIG. 2, and theeighth gear speed is established.

When the third engagement mechanism C3, the fifth engagement mechanismB2 and the sixth engagement mechanism B3 are engaged, the rotation speedof the sun gear Sb (the fourth element) of the second planetary gearmechanism 5 and the rotation speed of the ring gear Rc (the eighthelement) of the third planetary gear mechanism 6 become equal to “0”,the carrier Ca (the second element) of the first planetary gearmechanism 4 and the second coupling body Rb-Sc (the sixth-ninthelements) rotate at the same rotation speed. Thereby, the velocity lineof the three planetary gear mechanisms 4, 5 and 6 becomes a line denotedby “Rev” in FIG. 2, and the reverse gear speed is established.

The dotted velocity line in FIG. 2 denotes that each element of theother planetary gears in the three planetary gear mechanisms 4, 5 and 6is rotating in follow of the planetary gear transmitting the drivingforce.

FIG. 3 is an explanatory diagram illustrating relationships between eachof the mentioned gear speeds and each of the engagement states of thementioned engagement mechanisms of C1 to C3 and B1 to B3. The sign “O”denotes engagement. A gear ratio (rotation speed of the input shaft2/rotation speed of the output member 3) of each gear speed when thegear ratio i of the first planetary gear mechanism 4, the gear ratio jof the second planetary gear mechanism 5 and the gear ratio k of thethird planetary gear mechanism 6 are set to 1.666, 1.666 and 2.666,respectively, is represented in FIG. 3. Thereby, a common ratio (a ratiobetween gear ratios of each gear speed) becomes appropriate, and a ratiorange (the first speed ratio/ the eighth speed ratio) of 8 gear speedsdenoted in the common ratio column also becomes appropriate.

According to the automatic transmission of the first embodiment, thegear change can be performed to generate the 8 forward gear speeds, andfor each gear speed, there are three engagement mechanisms out of thesix engagement mechanisms of the first engagement mechanism C1 to thethird engagement mechanism C3 and the fourth engagement mechanism B1 tothe sixth engagement B3 engaged. Thereby, in each gear speed change, thefreed engagement mechanisms are three. In comparison with theconventional transmission in which the freed engagement mechanisms arefour, the friction loss caused by the freed engagement mechanisms isreduced, and consequently, the efficiency of the automatic transmissionis improved.

The automatic transmission of the first embodiment is described toperform gear change between the 8 forward gear speeds; however, it isacceptable to configure the automatic transmission to perform gearchange between 7 forward gear speeds by omitting either one of the gearspeeds. For example, the seventh gear speed in the first embodiment maybe omitted and the eighth gear speed is set as the seventh gear speed,the automatic transmission can perform gear change between 7 forwardgear speeds.

In the automatic transmission of the first embodiment, the firstplanetary gear mechanism 4 is configured as a single-pinion type;however, it is acceptable to configure the first planetary gearmechanism 4 to a double-pinion type. In this regard, for example, thefirst element may be set as the sun gear Sa, the second element may beset as the ring gear Ra and the third element may be set as the carrierCa.

In the automatic transmission of the first embodiment, the secondplanetary gear mechanism 5 is configured as a single-pinion type;however, as an automatic transmission of a second embodiment illustratedin FIG. 4, it is acceptable to configure the second planetary gearmechanism 5 to a double-pinion type. In this regard, for example, thefourth element may be set as the sun gear Sb, the fifth element may beset as the ring gear Rb and the sixth element may be set as the carrierCb.

In the automatic transmission of the first embodiment, the sixthengagement mechanism is configured as the third brake B3 and the 1-wayclutch F1; however, it is acceptable to omit the 1-way clutch F1 andleave only the third brake B3.

Moreover, as an automatic transmission of a third embodiment illustratedin FIG. 5, it is acceptable to configure the sixth engagement mechanismto a 2-way clutch F2 capable of switching between a state where thepositive rotation (forward rotation) of the ring gear Rc of the thirdplanetary gear mechanism 6 is allowed and the negative rotation (reverserotation) thereof is prevented and a state where the positive rotationof the ring gear Rc is prevented and the negative rotation thereof isallowed.

In this regard, similar to the case where the 1-way clutch F1 isdisposed, the control on the gear change between the first gear speedand the second gear speed can be improved. Moreover, the third brake B3,which is used in reverse gear change and has a relatively large volume,may be omitted to further reduce the friction loss, and consequently,the efficiency of the automatic transmission can be improved.

In the automatic transmission of the third embodiment illustrated inFIG. 5, the second engagement mechanism is comprised of a dog clutch D1(intermeshing mechanism). As clearly illustrated in FIG. 3, the secondengagement mechanism (equivalent to the second clutch C2 of the firstembodiment) is engaged in the range of low gear speeds from the firstgear speed to the fifth gear speed and released in the range of highgear speeds from the sixth gear speed to the eighth gear speed.

Thus, in the range of low gear speeds from the first to the fifth gearspeed where a torque difference between adjacent gear speeds is greaterin comparison with the range of high gear speeds, the switching betweenthe engaging states is not performed. The engaging states will beswitched only between the fifth gear speed and the sixth gear speedwhere the torque difference is relatively small. Thereby, the gearchange between the fifth gear speed and the sixth gear speed can beperformed smoothly.

Different from the wet multi-plate clutch engaged through friction,since the dog clutch D1 is engaged through mechanical intermeshing,there is no friction loss occurred. Therefore, compared with the secondengagement mechanism configured from the wet multi-plate clutch C2 inthe first embodiment, by configuring the second engagement mechanismfrom the dog clutch D1, it is possible to further reduce the frictionloss in the range of high gear speeds, improving the millage of thevehicle. In addition, it is acceptable to configure the dog clutch D1 asa synchromesh mechanism or the like having synchromesh functions.

An automatic transmission according to a fourth embodiment of thepresent invention is illustrated in FIG. 6. The automatic transmissionof the fourth embodiment is provided with a transmission case 1, aninput shaft 2 and an output member 3. The input shaft 2 is pivotallysupported inside the transmission case 1 and coupled with a drivingsource such as an engine (not shown). The output member 3 is comprisedof output gears disposed concentrically with the input shaft 2.Rotations of the output member 3 are transmitted to driving wheelsdisposed at both sides of a vehicle via a differential gear (not shown).

Further, a first planetary gear mechanism 4, a second planetary gearmechanism 5 and a third planetary gear mechanism 6 are disposedconcentrically with the input shaft 2 inside the transmission case 1.The first planetary gear mechanism 4 is of a single-pinion planetarygear mechanism comprised of a sun gear Sa, a ring gear Ra, and a carrierCa pivotally supporting a pinion Pa intermeshed with the sun gear Sa andthe ring gear Ra in such a way that the pinion Pa can rotate and revolvefreely.

Referring to the top section of a velocity diagram (a diagramillustrating rotation velocities of three elements of the sun gear, thecarrier and the ring gear by straight lines) for the first planetarygear mechanism 4 in FIG. 7, if the three elements composed of the sungear Sa, the carrier Ca and the ring gear Ra of the first planetary gearmechanism 4 are arranged from the left side in the order of distancesrelative to gear ratios in the velocity diagram, they are equivalent tothe first element, the second element and the third element,respectively.

Herein, when the gear ratio (number of teeth of the ring gear/number ofteeth of the sun gear) of the first planetary gear mechanism 4 issupposed to be “i”, the ratio of a distance between the sun gear Sa andthe carrier Ca to a distance between the carrier Ca and the ring gear Rais set to i:1. In the velocity diagram, the lower horizontal line andthe upper horizontal line indicate a rotational speed of “0” and arotational speed of “1” (equal to that of the input shaft 2),respectively.

The second planetary gear mechanism 5 is of a single-pinion planetarygear mechanism comprised of a sun gear Sb, a ring gear Rb and a carrierCb pivotally supporting a pinion Pb intermeshed with the sun gear Sb andthe ring gear Rb in such a way that the pinion Pb can rotate and revolvefreely.

Referring to the middle section of a velocity diagram for the secondplanetary gear mechanism 5 in FIG. 7, if the three elements composed ofthe sun gear Sb, the carrier Cb and the ring gear Rb of the secondplanetary gear mechanism 5 are arranged from the left side in the orderof distances relative to gear ratios in the velocity diagram, they areequivalent to the fourth element, the fifth element and the sixthelement, respectively. When the gear ratio of the second planetary gearmechanism 5 is supposed to be “j”, the ratio of a distance between thesun gear Sb and the carrier Cb to a distance between the carrier Cb andthe ring gear Rb is set to j:1.

The third planetary gear mechanism 6 is of a single-pinion planetarygear mechanism comprised of a sun gear Sc, a ring gear Rc, and a carrierCc pivotally supporting a pinion Pc intermeshed with the sun gear Sc andthe ring gear Rc in such a way that the pinion Pc can rotate and revolvefreely.

Referring to the bottom section of a velocity diagram for the thirdplanetary gear mechanism 6 in FIG. 7, if the three elements composed ofthe ring gear Rc, the carrier Cc and the sun gear Sc of the thirdplanetary gear mechanism 6 are arranged from the left side in the orderof distances relative to gear ratios in the velocity diagram, they areequivalent to the seventh element, the eighth element and the ninthelement, respectively. When the gear ratio of the third planetary gearmechanism 6 is supposed to be “k”, the ratio of a distance between thesun gear Sc and the carrier Cc to a distance between the carrier Cc andthe ring gear Rc is set to k:1.

The automatic transmission of the fourth embodiment is provided with aparallel gear 7 disposed adjacent to the ring gear Rc (the seventhelement) of the third planetary gear mechanism 6 and intermeshed withthe pinion Pc of the third planetary gear mechanism 6. The parallel gear7 is configured to have a ring shape with the same inner diameter andthe same number of teeth on the inner circumference as the ring gear Rc(the seventh element) and to rotate at the same rotation speed as thering gear Rc (the seventh element) of the third planetary gear mechanism6.

The sun gear Sa (the first element) of the first planetary gearmechanism 4 is coupled with the input shaft 2. The parallel gear 7disposed in parallel to the ring gear Rc (the seventh element) of thethird planetary gear mechanism 6 is coupled with the output member 3equivalent to the output gear.

The ring gear Ra (the third element) of the first planetary gearmechanism 4 and the carrier Cb (the fifth element) of the secondplanetary gear mechanism 5 are coupled with each other to form a firstcoupling body Ra-Cb. The ring gear Rb (the sixth element) of the secondplanetary gear mechanism 5 and the sun gear Sc (the ninth element) ofthe third planetary gear mechanism 6 are coupled with each other to forma second coupling body Rb-Sc.

In the automatic transmission of the fourth embodiment, a total numberof 7 rotation bodies are constituted in the three planetary gearmechanisms 4, 5 and 6, specifically, the sun gear Sa (the first element)and the carrier Ca (the second element) of the first planetary gearmechanism 4, the first coupling body Ra-Cb (the third-fifth elements),the sun gear Sb (the fourth element) of the second planetary gearmechanism 5, the second coupling body Rb-Sc (the sixth-ninth elements),the ring gear Rc (the seventh element) and the carrier Cc (the eighthelement) of the third planetary gear mechanism 6.

The automatic transmission of the fourth embodiment is provided withengagement mechanisms comprised of wet multi-plate clutches,specifically, a first clutch C1 equivalent to a first engagementmechanism coupled with the input shaft 2 and the carrier Cc (the eighthelement) of the third planetary gear mechanism 6 releasably, a secondclutch C2 equivalent to a second engagement mechanism coupled with thecarrier Ca (the second element) of the first planetary gear mechanism 4and the ring gear Rc (the seventh element) of the third planetary gearmechanism 6 releasably, and a third clutch C3 equivalent to a thirdengagement mechanism coupled with the carrier Ca (the second element) ofthe first planetary gear mechanism 4 and the second coupling body Rb-Screleasably.

The clutch C3 equivalent to the third engagement mechanism is disposedat an outer position of the second clutch C2 equivalent to the secondengagement mechanism in the radial direction thereof and overlapped withthe second clutch C2 in the axial direction of the input shaft 2 toshorten the shaft length of the automatic transmission.

The automatic transmission of the fourth embodiment is provided withengagement mechanisms comprised of wet multi-plate brakes, specifically,a first brake B1 equivalent to a fourth engagement mechanism fixing thefirst coupling body Ra-Cb (the third-fifth elements) to the transmissioncase 1 releasably, a second brake B2 equivalent to a fifth engagementmechanism fixing the sun gear Sb of the second planetary gear mechanism5 to the transmission case 1 releasably, and a third brake B3 fixing thecarrier Cc (the eighth element) of the third planetary gear mechanism 6to the transmission case 1 releasably.

A 1-way clutch F1 is disposed in parallel with the third brake B3 insidethe transmission case 1, allowing the carrier Cc (the eighth element) ofthe third planetary gear mechanism 6 to rotate positively (forwardrotation) and preventing it from rotating negatively (reverse rotation).

The third brake B3 and the 1-way clutch F1 are disposed at an outerposition of the ring gear Rc (the seventh element) of the thirdplanetary gear mechanism 6 in the radial direction thereof The thirdbrake B3 and the 1-way clutch F1 in the automatic transmission of thefourth embodiment constitute the sixth engagement mechanism of thepresent invention.

When the second clutch C2 (the second engagement mechanism) and thesecond brake B2 (the fifth engagement mechanism) are engaged in theautomatic transmission of the fourth embodiment, the carrier Ca (thesecond element) of the first planetary gear mechanism 4 and the ringgear Rc (the seventh element) of the third planetary gear mechanism 6rotate at the same rotation speed, the rotation speed of the sun gear Sb(the fourth element) of the second planetary gear mechanism 5 becomesequal to “0”, and the rotation speed of the carrier Cc (the eighthelement) of the third planetary gear mechanism 6 becomes equal to “0”due to the function of the 1-way clutch F1. Thereby, the velocity lineof the three planetary gear mechanisms 4, 5 and 6 becomes a line denotedby “1st” in FIG. 7, and the first gear speed is established.

At this time, the third brake B3 is released, however, since therotation speed of the carrier Cc (the eighth element) of the thirdplanetary gear mechanism 6 becomes equal to “0” due to the function ofthe 1-way clutch F1, no friction loss will be happened in the thirdbrake B3. Moreover, the disposition of the 1-way clutch F1 makes itunnecessary to supply pressured oils to the third brake B3 and to stopsupplying pressured oils thereto when the gear change is made betweenthe first gear speed and the second gear speed, which improves thecontrol on gear change between the first gear speed and the second gearspeed.

In addition to the second clutch C2 (the second engagement mechanism)and the second brake B2 (the fifth engagement mechanism), when the thirdbrake B3 is further engaged, the first gear speed is established withthe engine braking in action.

When the second clutch C2, the first brake B1 and the second brake B2are engaged, both the rotation speed of the first coupling body Ra-Cb(the third-fifth elements) and the rotation speed of the sun gear Sb(the fourth element) of the second planetary gear mechanism 5 becomeequal to “0”, the three elements of the second planetary gear mechanism5 are locked in a state where relative rotations are impossible,therefore, the rotation speed of the second coupling body Rb-Sc (thesixth-ninth elements) also becomes equal to “0”.

Thereby, the carrier Ca (the second element) of the first planetary gearmechanism 4 and the ring gear Rc (the seventh element) of the thirdplanetary gear mechanism 6 rotate at the same rotation speed, thevelocity line of the three planetary gear mechanisms 4, 5 and 6 becomesa line denoted by “2nd” in FIG. 7, and the second gear speed isestablished.

When the second clutch C2, the third clutch C3 and the second brake B2are engaged, the carrier Ca (the second element) of the first planetarygear mechanism 4, the second coupling body Rb-Sc (the sixth-ninthelements) and the ring gear Rc (the seventh element) of the thirdplanetary gear mechanism 6 rotate at the same rotation speed, the threeelements, namely sun gear Sc, the carrier Cc and the ring gear Rc of thethird planetary gear mechanism 6 are locked in a state where relativerotations are impossible. Thereby, the velocity line of the threeplanetary gear mechanisms 4, 5 and 6 becomes a line denoted by “3rd” inFIG. 7, and the third gear speed is established.

When the first clutch C1, the second clutch C2 and the second brake B2are engaged, the rotation speed of the sun gear Sa (the first element)of the first planetary gear mechanism 4 and the rotation speed of thecarrier Cc (the eighth element) of the third planetary gear mechanism 6become equal to “1”, and the carrier Ca (the second element) of thefirst planetary gear mechanism 4 and the ring gear Rc (the seventhelement) of the third planetary gear mechanism 6 rotate at the samerotation speed. Thereby, the velocity line of the three planetary gearmechanisms 4, 5 and 6 becomes a line denoted by “4th” in FIG. 7, and thefourth gear speed is established.

When the first clutch C1, the second clutch C2 and the third clutch C3are engaged, the rotation speed of the sun gear Sa (the first element)of the first planetary gear mechanism 4 and the rotation speed of thecarrier Cc (the eighth element) of the third planetary gear mechanism 6become equal to “1”, the three elements, namely the sun gear Sc, thecarrier Cc and the ring gear Rc of the third planetary gear mechanism 6are locked in a state where relative rotations are impossible. Thereby,the fifth gear speed is established at “1” which is also the rotationspeed of the carrier Cc of the third planetary gear mechanism 6 coupledwith the output member 3.

When the first clutch C1, the third clutch C3 and the second brake B2are engaged, the rotation speed of the sun gear Sa (the first element)of the first planetary gear mechanism 4 and the rotation speed of thecarrier Cc (the eighth element) of the third planetary gear mechanism 6become equal to “1”, the rotation speed of the sun gear Sb (the fourthelement) of the second planetary gear mechanism 5 becomes equal to “0”,and the carrier Ca (the second element) of the first planetary gearmechanism 4 and the second coupling body Rb-Sc (the sixth-ninthelements) rotate at the same rotation speed. Thereby, the velocity lineof the three planetary gear mechanisms 4, 5 and 6 becomes a line denotedby “6th” in FIG. 7, and the sixth gear speed is established.

When the first clutch C1, the third clutch C3 and the first brake B1 areengaged, the rotation speed of the sun gear Sa (the first element) ofthe first planetary gear mechanism 4 and the rotation speed of thecarrier Cc (the eighth element) of the third planetary gear mechanism 6become equal to “1”, the rotation speed of the first coupling body Ra-Cb(the third-fifth elements) becomes equal to “0”, and the carrier Ca (thesecond element) of the first planetary gear mechanism 4 and the secondcoupling body Rb-Sc (the sixth-ninth elements) rotate at the samerotation speed. Thereby, the velocity line of the three planetary gearmechanisms 4, 5 and 6 becomes a line denoted by “7th” in FIG. 7, and theseventh gear speed is established.

When the first clutch C1, the first brake B1 and the second brake B2 areengaged, the rotation speed of the sun gear Sa (the first element) ofthe first planetary gear mechanism 4 and the rotation speed of thecarrier Cc (the eighth element) of the third planetary gear mechanism 6become equal to “1”, the rotation speed of the sun gear Sb (the fourthelement) of the second planetary gear mechanism 5 and the rotation speedof the first coupling body Ra-Cb (the third-fifth elements) become equalto “0”.

Thereby, the three elements, namely the sun gear Sb, the carrier Cb andthe ring gear Rb of the second planetary gear mechanism 5 are locked ina state where relative rotations are impossible, the rotation speed ofthe second coupling body Rb-Sc (the sixth-ninth elements) becomes equalto “0”. Thereby, the velocity line of the three planetary gearmechanisms 4, 5 and 6 becomes a line denoted by “8th” in FIG. 7, and theeighth gear speed is established.

When the third clutch C3, the second brake B2 and the third brake B3 areengaged, the rotation speed of the sun gear Sb (the fourth element) ofthe second planetary gear mechanism 5 and the rotation speed of thecarrier Cc (the eighth element) of the third planetary gear mechanism 6become equal to “0”, the carrier Ca (the second element) of the firstplanetary gear mechanism 4 and the second coupling body Rb-Sc (thesixth-ninth elements) rotate at the same rotation speed. Thereby, thevelocity line of the three planetary gear mechanisms 4, 5 and 6 becomesa line denoted by “Rev” in FIG. 7, and the reverse gear speed isestablished.

The dotted velocity line in FIG. 7 denotes that each element of theother planetary gears in the three planetary gear mechanisms 4, 5 and 6is rotating in follow of the planetary gear transmitting the drivingforce.

FIG. 8 is an explanatory diagram illustrating relationships between eachof the mentioned gear speeds and each of the engagement states of thementioned engagement mechanisms of C1 to C3 and B1 to B3. The sign “O”denotes engagement. A gear ratio (rotation speed of the input shaft2/rotation speed of the output member 3) of each gear speed when thegear ratio i of the first planetary gear mechanism 4, the gear ratio jof the second planetary gear mechanism 5 and the gear ratio k of thethird planetary gear mechanism 6 are set to 1.666, 1.666 and 1.666,respectively, is represented in FIG. 8. Thereby, a common ratio (a ratiobetween gear ratios of each gear speed) becomes appropriate, and a ratiorange (the first speed ratio/the eighth speed ratio) of 8 gear speedsdenoted in the common ratio column also becomes appropriate.

According to the automatic transmission of the fourth embodiment, thegear change can be performed to generate the 8 forward gear speeds, andfor each gear speed, there are three engagement mechanisms out of thesix engagement mechanisms of the first engagement mechanism C1 to thethird engagement mechanism C3 and the fourth engagement mechanism B1 tothe sixth engagement B3 engaged. Thereby, in each gear speed change, thefreed engagement mechanisms are three. In comparison with theconventional transmission in which the freed engagement mechanisms arefour, the friction loss caused by the freed engagement mechanisms isreduced, and consequently, the efficiency of the automatic transmissionis improved.

Although structurally it is impossible to couple directly the ring gearRc (the seventh element) of the third planetary gear mechanism 6 withthe output member 3, by coupling the output member 3 with the parallelgear 7 which rotates at the same rotation speed with the ring gear Rc(the seventh element), the rotation speed of the ring gear Rc (theseventh element) of the third planetary gear mechanism 6 can be outputout through the output member 3.

The automatic transmission of the fourth embodiment is described toperform gear change between the 8 forward gear speeds; however, it isacceptable to configure the automatic transmission to perform gearchange between 7 forward gear speeds by omitting either one of the gearspeeds. For example, the seventh gear speed in the fourth embodiment maybe omitted and the eighth gear speed is set as the seventh gear speed,the automatic transmission can perform gear change between 7 forwardgear speeds.

In the automatic transmission of the fourth embodiment, the firstplanetary gear mechanism 4 is configured as a single-pinion type;however, it is acceptable to configure the first planetary gearmechanism 4 to a double-pinion type. In this regard, for example, thefirst element may be set as the sun gear Sa, the second element may beset as the ring gear Ra and the third element may be set as the carrierCa.

In the automatic transmission of the fourth embodiment, the secondplanetary gear mechanism 5 is configured as a single-pinion type;however, as an automatic transmission of a fifth embodiment illustratedin FIG. 9, it is acceptable to configure the second planetary gearmechanism 5 to a double-pinion type. In this regard, for example, thefourth element may be set as the sun gear Sb, the fifth element may beset as the ring gear Rb and the sixth element may be set as the carrierCb.

In the automatic transmission of the fourth embodiment, the sixthengagement mechanism is configured as the third brake B3 and the 1-wayclutch F1; however, it is acceptable to omit the 1-way clutch F1 andleave only the third brake B3.

Moreover, as the automatic transmission of the fifth embodimentillustrated in FIG. 9, it is acceptable to configure the sixthengagement mechanism to a 2-way clutch F2 capable of switching between astate where the positive rotation (forward rotation) of the ring gear Rcof the third planetary gear mechanism 6 is allowed and the negativerotation (reverse rotation) thereof is prevented and a state where thepositive rotation of the ring gear Rc is prevented and the negativerotation thereof is allowed.

In this regard, similar to the case where the 1-way clutch F1 isdisposed, the control on the gear change between the first gear speedand the second gear speed can be improved. Moreover, the third brake B3,which is used in reverse gear change and has a relatively large volume,may be omitted to further reduce the friction loss, and consequently,the efficiency of the automatic transmission can be improved.

In the automatic transmission of the fifth embodiment illustrated inFIG. 9, the second engagement mechanism is comprised of a dog clutch D1(intermeshing mechanism). As clearly illustrated in FIG. 8, the secondengagement mechanism (equivalent to the second clutch C2 of the fourthembodiment) is engaged in the range of low gear speeds from the firstgear speed to the fifth gear speed and released in the range of highgear speeds from the sixth gear speed to the eighth gear speed.

Thus, in the range of low gear speeds from the first to the fifth gearspeed where a torque difference between adjacent gear speeds is greaterin comparison with the range of high gear speeds, the switching betweenthe engaging states is not performed. The engaging states will beswitched only between the fifth gear speed and the sixth gear speedwhere the torque difference is relatively small. Thereby, the gearchange between the fifth gear speed and the sixth gear speed can beperformed smoothly.

Different from the wet multi-plate clutch engaged through friction,since the dog clutch D1 is engaged through mechanical intermeshing,there is no friction loss occurred. Therefore, compared with the secondengagement mechanism configured from the wet multi-plate clutch C2 inthe fourth embodiment, by configuring the second engagement mechanismfrom the dog clutch D1, it is possible to further reduce the frictionloss in the range of high gear speeds, improving the millage of thevehicle. In addition, it is acceptable to configure the dog clutch D1 asa synchromesh mechanism or the like having synchromesh functions.

In the present embodiment, the parallel gear 7 is configured to have aring shape with the same inner diameter and the same number of teeth onthe inner circumference as the ring gear Rc (the seventh element) of thethird planetary gear mechanism 6; however, it is not limited thereto,the parallel gear 7 may be configured to have an inner diameter and thenumber of teeth different from the ring gear Rc (the seventh element).

In this case, for example, the pinion Pc of the third planetary gearmechanism 6 is configured as a stepped pinion with a smaller diameterportion and a greater diameter portion, the smaller diameter portion orthe greater diameter portion of the pinion Pc is configured to beintermeshed with the sun gear Sc and the ring gear Rc and the other partof the pinion Pc is configured to be intermeshed with the parallel gear7.

If the number of teeth of the parallel gear 7 divided by the number ofteeth of the sun gear Sc is set as “m” and the number of teeth of thegreater diameter portion or the smaller diameter portion of the pinionPc intermeshed with the parallel gear 7 divided by the number of teethof the other part of the pinion Pc intermeshed with the sun gear Sc isset as “n”, when the distance between the sun gear Sc and the carrier Ccis set as “1” in the velocity diagram, the parallel gear 7 is disposedat a position left to the carrier Cc so that the distance between theparallel gear 7 and the carrier Cc satisfies n/m, which is differentfrom the position of the ring gear Rc in the velocity diagram. Thereby,the parallel gear 7 rotates at a rotation speed different from the ringgear Rc.

According to the mentioned configuration, by changing the number ofteeth of the parallel gear 7, the gear ratio of each gear speed can beset with more freedom. Furthermore, by configuring the parallel gear 7into a small diameter, it will be more flexible to dispose the othercomponents in the automatic transmission.

However, if the output member 3 is disposed between the second brake B2and the third brake B3 as described in the fourth embodiment, when theautomatic transmission is used in a FR layout vehicle, it is necessaryto dispose a counter shaft having a gear intermeshed with the outputmember 3 and the driving force is transmitted to the rear wheels at bothsides through a propeller shaft coupled with the counter shaft. Theincrement on the number of components such as the counter shaft makes itimpossible to miniaturize the automatic transmission.

In this regard, in the automatic transmission of the fifth embodiment asillustrated in FIG. 9, the third brake B3 is coupled to the carrier Cc(the eighth element) of the third planetary gear mechanism 6 at the sideof the driving source, and the first brake B1 to the third brake B3 aredisposed closer to the driving source than the parallel gear 7, theoutput member 3 comprised of an output shaft disposed coaxially with theinput shaft 2 can be coupled with the parallel gear 7 without beinghindered by the third brake B3. Therefore, the output member 3 can becoupled with the propeller shaft without using the counter shaft, whichmakes it possible to miniaturize the automatic transmission to be usedin a FR layout vehicle.

What is claimed is:
 1. An automatic transmission which changes rotationsof an input shaft into multiple gear speeds transmitted to an outputmember via a plurality of planetary gear mechanisms disposed in atransmission case, wherein the plurality of planetary gear mechanismsincludes three planetary gear mechanisms of a first planetary gearmechanism to a third planetary gear mechanism, the third planetary gearmechanism is of a single-pinion planetary gear mechanism comprised ofthe sun gear, the ring gear, and the carrier pivotally supporting apinion intermeshed with the sun gear and the ring gear in such a waythat the pinion can rotate and revolve freely, three elements comprisedof a sun gear, a carrier and a ring gear of the first planetary gearmechanism are set as a first element, a second element and a thirdelement, respectively, in the order of distances relative to gear ratiosin a velocity diagram, three elements comprised of a sun gear, a carrierand a ring gear of the second planetary gear mechanism are set as afourth element, a fifth element and a sixth element, respectively, inthe order of distances relative to gear ratios in the velocity diagram,three elements comprised of a sun gear, a carrier and a ring gear of thethird planetary gear mechanism are set as a seventh element, an eighthelement and a ninth element, respectively, in the order of distancesrelative to gear ratios in the velocity diagram, a parallel gear isdisposed to be adjacent to the seventh element and intermeshed with thepinion of the third planetary gear mechanism, the first element iscoupled with the input shaft, the parallel gear is coupled with theoutput member, the third element and the fifth element are coupled toform a first coupling body, the sixth element and the ninth element arecoupled to form a second coupling body, a first engagement mechanismcouples the input shaft with the eighth element releasably, a secondengagement mechanism couples the second element with the seventh elementreleasably, a third engagement mechanism couples the second element withthe second coupling body releasably, a fourth engagement mechanism fixesthe first coupling body to the transmission case releasably, a fifthengagement mechanism fixes the fourth element to the transmission casereleasably, and a sixth engagement mechanism fixes the eighth element tothe transmission case releasably.
 2. The automatic transmissionaccording to claim 1, wherein both the first planetary gear mechanismand the second planetary gear mechanism are of a single-pinion planetarygear mechanism comprised of the sun gear, the ring gear, and the carrierpivotally supporting a pinion intermeshed with the sun gear and the ringgear in such a way that the pinion can rotate and revolve freely.
 3. Theautomatic transmission according to claim 1, wherein the sixthengagement mechanism is a one-way clutch or a two-way clutch.
 4. Theautomatic transmission according to claim 1, wherein the secondengagement mechanism is an intermeshing mechanism.
 5. The automatictransmission according to claim 1, wherein the fourth to the sixthengagement mechanisms are disposed closer to a driving source fordriving the input shaft to rotate than to the parallel gear, and theoutput member is an output shaft disposed coaxially with the inputshaft.